Simplified mechanism for a scotch yoke actuator

ABSTRACT

An actuator converts linear motion into rotational motion. The actuator includes a power piston having a longitudinal axis. The power piston is configured to move back and forth along the longitudinal axis. A slider bearing is disposed within a cross-bore opening in the power piston. The cross-bore opening has a slider bearing axis orthogonal to the longitudinal axis. The slider bearing is configured to move back and forth within the cross-bore opening. An output shaft has a central axis orthogonal to both the longitudinal and the slider bearing axes. The output shaft includes an eccentric shaft pin inserted through a slot in the power piston and disposed within a cavity of the slider bearing. The eccentric shaft pin is offset from the central axis of the output shaft. The reciprocating movement of the power piston causes the slider bearing to move back and forth rotating the output shaft.

FIELD OF THE INVENTION

This invention generally relates to air valve actuators for jet engines,and more particularly to Scotch Yoke Actuators.

BACKGROUND OF THE INVENTION

Jet engines employ air valves for a variety of reasons including, butnot limited to, passenger compartment cooling, turbine clearance, andcompressor bleed. Generally, these air valves require some type ofrotary actuation. Different types of actuator mechanizations have beendeveloped for these applications including, but not limited to, rotaryvane actuators (RVA), linear crank sliders, Scotch yoke actuators,reverse Scotch yoke actuators, rotary piston actuators (RPA), and rackand pinion actuators.

The most common mechanism for converting linear force into rotationaltorque in air valves used on aircraft turbine engines is the linearcrank slider. The linear crank slider has several disadvantages, chiefamong them: 1) Large overhanging mass due to length needed to implementthe mechanization which adds to the size, weight, and cost of thesystem; 2) Linear seal and scraper at fluid/air boundary in harshenvironment which adversely affect reliability; and 3) Linkage andassociated bearings exposed to harsh environment which adversely affectreliability.

It would therefore be desirable to have an air valve actuation systemthat addresses the shortcoming recited above. Embodiments of the presentinvention provide such an air valve actuation system. These and otheradvantages of the invention, as well as additional inventive features,will be apparent from the description of the invention provided herein.

BRIEF SUMMARY OF THE INVENTION

In one aspect, embodiments of the invention provide an actuator forconverting linear motion into rotational motion is provided. Theactuator includes a cylinder having a longitudinal axis and a powerpiston configured to move in a reciprocating manner back and forthwithin the cylinder along the longitudinal axis. A slider bearing isdisposed within a cross-bore opening in the power piston. The cross-boreopening has a slider bearing axis that is orthogonal to the longitudinalaxis. The slider bearing is configured to move back and forth within thecross-bore opening. An output shaft has a central axis that isorthogonal to both the longitudinal axis and the slider bearing axis.The output shaft includes an eccentric shaft pin inserted through a slotin the power piston and disposed within a cavity of the slider bearing.The eccentric shaft pin is offset from the central axis of the outputshaft. The reciprocating movement of the power piston causes the sliderbearing to move back and forth rotating the output shaft.

In certain embodiments of the invention, the slider bearing is made froma self-lubricating material. The claimed actuator may also include afirst radial bushing assembled onto the output shaft and adjacent an endof the output shaft which includes the eccentric shaft pin. In someembodiments, the first radial bushing is an annular radial bushing.

In other embodiments, the claimed actuator includes a first shaft sealassembled onto the output shaft to prevent a leakage of fluid from thecylinder to the output shaft. Moreover, the claimed actuator may furtherinclude a second radial bushing and second shaft seal each assembledonto the output shaft proximate the first radial bushing and first shaftseal. In a particular embodiment, the first shaft seal is an annularseal.

Further, in certain embodiments, the output shaft and eccentric shaftpin are monolithic being constructed as a single component.Additionally, the output shaft and eccentric shaft pin may include agimbaling connection to accommodate deflection of the output shaft.

In another aspect, embodiments of the invention provide an air valvethat includes a valve housing, and a moveable valve member disposedwithin the valve housing. The air valve also includes a valve actuatorcoupled to the valve member. The valve actuator includes a power pistonhaving a longitudinal axis, the power piston configured to move in areciprocating manner back and forth within the cylinder along thelongitudinal axis. A slider bearing is disposed within a cross-boreopening in the power piston. The cross-bore opening has a slider bearingaxis that is orthogonal to the longitudinal axis. The slider bearing isconfigured to move back and forth within the cross-bore opening. Anoutput shaft has a central axis that is orthogonal to both thelongitudinal axis and the slider bearing axis. The output shaft includesan eccentric shaft pin inserted through a slot in the power piston anddisposed within a cavity of the slider bearing. The eccentric shaft pinis offset from the central axis of the output shaft. Also, thereciprocating movement of the power piston causes the slider bearing tomove back and forth rotating the output shaft.

In a particular embodiment, the output shaft extends through the valvehousing and controls a rotational movement of the valve member. In amore particular embodiment, the air valve is a butterfly valve, and thevalve member is a disk which the output shaft rotates to control a flowof fluid through the valve housing. In an alternate embodiment, the airvalve is a ball valve, and the valve member is a ball which the outputshaft rotates to control a flow of fluid through the valve housing.

In certain embodiments, the air valve includes an electrohydraulic servovalve (EHSV) coupled to a cylinder which houses the power piston, theelectrohydraulic servo valve configured to control linear movement ofthe power piston. In a more particular embodiment, the air valveincludes a linear variable differential transformer disposed in thecylinder, the linear variable differential transformer configured toprovide positional information for the power piston.

In some embodiments of the air valve, a first radial bushing isassembled onto the output shaft. The first radial bushing is adjacent anend of the output shaft which includes the eccentric shaft pin. Anembodiment also includes a first shaft seal assembled onto the outputshaft to prevent a leakage of fluid from the cylinder to the outputshaft. Other embodiments include a second radial bushing and secondshaft seal each assembled onto the output shaft proximate the firstradial bushing and first shaft seal.

A further embodiment of the air valve includes a stepper motor having arotor. A cam is operatively connected to the stepper motor. The camposition rotates in response to rotation of the rotor. The cam isdisposed within a control valve body having an inlet port, a rod port, ahead port and at least one drain port. A control piston is positionedwithin the control valve body and has a nozzle positioned on a firstside of the cam in close proximity to a surface of the cam, and ismovable between a null position and flow positions. The nozzle has aflow path leading from a first end of the control piston. The controlpiston further includes a means for applying a force on a second end ofthe control piston. The control piston moves as a result of a pressureimbalance at the ends of the control piston occurring in response to achange in position of the cam. An actuator piston is operably coupled tothe cam. The actuator piston has a first side and a second side. Thefirst side is in fluid communication with the head port and the secondside is in fluid communication with the rod port. The cam moves inresponse to movement of the actuator piston. The actuator pistongenerates the reciprocating movement of the power piston which, in turn,causes the slider bearing to move back and forth rotating the outputshaft.

Further embodiments of the air valve include those in which the cam isconnected to a gear shaft which rotates the cam, the actuator furthercomprising a gearbox connected between the rotor and the gear shaft. Incertain embodiments, the translation of the control piston opens one ofthe head port and rod port to supply and the other of the head port androd to drain, thereby causing the actuator piston and rack to stroke,wherein translation of the control piston towards a first end of thecontrol valve body opens the head port to supply and the rod port todrain, and wherein translation of the control piston towards a secondend of the control valve body opens the head port to drain and the rodport to supply, and wherein the actuator piston rotates the cam andtranslates the control piston to a mechanical null position in responseto the rack stroking.

In particular embodiments, the means for applying a force on a secondend of the control piston is a spring in operative contact with thesecond end of the control piston, wherein the second end of the controlpiston is in fluid communication with the drain port. In otherembodiments, the means for applying a force on a second end of thecontrol piston is a double diameter end portion having an end thereof influid communication with the inlet port, and wherein a hydraulicpressure at the end of the double diameter varies in the same manner asa hydraulic pressure at the first end of the control piston.

In yet another aspect, embodiments of the invention provide an actuatorfor converting linear motion into rotational motion. The actuatorincludes a power piston configured to move in a reciprocating mannerback and forth within the cylinder along a longitudinal axis. A sliderbearing is disposed within a cross-bore opening in the power piston. Thecross-bore opening has a slider bearing axis that is orthogonal to thelongitudinal axis. The slider bearing is configured to move back andforth within the cross-bore opening. An output shaft has a central axisthat is orthogonal to both the longitudinal axis and the slider bearingaxis. The output shaft includes an eccentric shaft pin inserted througha slot in the power piston and disposed within a cavity of the sliderbearing. The eccentric shaft pin is offset from the central axis of theoutput shaft. The actuator includes a stepper motor having a rotor. Acam is operatively connected to the stepper motor. The cam positionrotates in response to rotation of the rotor. The cam is disposed withina control valve body having an inlet port, a rod port, a head port andat least one drain port. A control piston is positioned within thecontrol valve body and has a nozzle positioned on a first side of thecam in close proximity to a surface of the cam, and is movable between anull position and flow positions. The nozzle has a flow path leadingfrom a first end of the control piston. The control piston furtherincludes a means for applying a force on a second end of the controlpiston. The control piston moves as a result of a pressure imbalance atthe ends of the control piston occurring in response to a change inposition of the cam. An actuator piston is operably coupled to the cam.The actuator piston has a first side and a second side. The first sideis in fluid communication with the head port and the second side is influid communication with the rod port. The cam moves in response tomovement of the actuator piston. The actuator piston generates thereciprocating movement of the power piston which, in turn, causes theslider bearing to move back and forth rotating the output shaft.

In a particular embodiment, the slider bearing is made from aself-lubricating material. In a further embodiment, the output shaft andeccentric shaft pin are monolithic being constructed as a singlecomponent, and wherein the output shaft and eccentric shaft pin includea gimbaling connection to accommodate deflection of the output shaft.Embodiments of the invention include those in which the cam is connectedto a gear shaft which rotates the cam, the actuator further comprising agearbox connected between the rotor and the gear shaft.

In a particular embodiment, the translation of the control piston opensone of the head port and rod port to supply and the other of the headport and rod to drain, thereby causing the actuator piston and rack tostroke. In a further embodiment, translation of the control pistontowards a first end of the control valve body opens the head port tosupply and the rod port to drain, and translation of the control pistontowards a second end of the control valve body opens the head port todrain and the rod port to supply.

In some embodiments, the actuator piston rotates the cam and translatesthe control piston to a mechanical null position in response to the rackstroking. The means for applying a force on a second end of the controlpiston may include a spring in operative contact with the second end ofthe control piston, wherein the second end of the control piston is influid communication with the drain port. Alternatively, the means forapplying a force on a second end of the control piston may include adouble diameter end portion having an end thereof in fluid communicationwith the inlet port, and wherein a hydraulic pressure at the end of thedouble diameter varies in the same manner as a hydraulic pressure at thefirst end of the control piston.

Embodiments of stepper-motor-driven actuators are disclosed in U.S. Pat.No. 7,963,185, issued to Spickard. This patent reference is herebyincorporated in its entirety.

Other aspects, objectives and advantages of the invention will becomemore apparent from the following detailed description when taken inconjunction with the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

The accompanying drawings incorporated in and forming a part of thespecification illustrate several aspects of the present invention and,together with the description, serve to explain the principles of theinvention. In the drawings:

FIG. 1 is a perspective view of a portion of a scotch yoke actuator,constructed in accordance with an embodiment of the invention;

FIG. 2 is a perspective view of a scotch yoke actuator of FIG. 1 showinga cross-section of the scotch yoke actuator;

FIG. 3 is a perspective cut-away view of a scotch yoke actuator,constructed in accordance with an embodiment of the invention;

FIG. 4 is a perspective view of an air valve incorporating the scotchyoke actuator according to an embodiment of the invention;

FIG. 5 is another perspective view of the air valve of FIG. 4incorporating the scotch yoke actuator;

FIGS. 6A and 6B show alternate embodiments of an air valve incorporatingthe scotch yoke actuator, according to an embodiment of the invention;

FIG. 7 is a perspective view of a stepper-motor-driven scotch yokeactuator, according to an embodiment of the invention;

FIG. 8 is a cross-sectional view of the actuator system in accordancewith the teachings of the present invention;

FIG. 9 is a partial cross-sectional view of the actuator system of FIG.8 with the stepper motor shown as a separate component for clarity andthe control piston at a centered position;

FIG. 10 is a schematic view of the actuator system of FIG. 9illustrating the cam-rack interaction;

FIG. 11 is a partial cross-sectional view of the actuator system of FIG.9 with the control piston at a position such that flow drives theactuator in the retract direction with the actuator against the retractstop;

FIG. 12 is a partial cross-sectional view of the actuator system of FIG.2 with the control piston at a position such that flow drives theactuator in the extend direction with the actuator against the extendstop;

FIG. 13 is a partial cross-sectional view of a single nozzle embodimentof the actuator system constructed in accordance with the teachings ofthe present invention; and

FIG. 14 is a partial cross-sectional view of another single nozzleembodiment of the actuator system constructed in accordance with theteachings of the present invention.

While the invention will be described in connection with certainpreferred embodiments, there is no intent to limit it to thoseembodiments. On the contrary, the intent is to cover all alternatives,modifications and equivalents as included within the spirit and scope ofthe invention as defined by the appended claims.

DETAILED DESCRIPTION OF THE INVENTION

FIG. 1 is a perspective view of a portion of a scotch yoke actuator 100,constructed in accordance with an embodiment of the invention. FIG. 2 isa perspective view of a scotch yoke actuator of FIG. 1 showing across-section of the scotch yoke actuator. The scotch yoke actuator 100includes a power piston 102 configured to move linearly in areciprocating motion in the direction shown by the arrow in FIG. 1. Inthe embodiments shown, the power piston 102 is cylindrical, though otherpiston configurations are envisioned. The power piston 102 has alongitudinal axis 104 and a cross-bore opening 106. Within thecross-bore opening 106, there is a slider bearing 108. The sliderbearing 108 has a slider bearing axis 110. The slider bearing axis 110is orthogonal to the longitudinal axis 104 of the power piston 102. Inparticular embodiments, the slider bearing 108 is made from aself-lubricating material.

In the embodiment of FIGS. 1 and 2, an output shaft 112 is positionedbelow the power piston 102 substantially at a right angle to the powerpiston 102. The output shaft 112 has a central axis 114 that isorthogonal to both the longitudinal axis 104 and the slider bearing axis110. At one end of the output shaft 112, there is an eccentric shaft pin116. The eccentric shaft pin 116 is offset from the central axis 114positioned proximate an outer diameter of the output shaft 112. Theeccentric shaft pin 116 extends through a slot 118 in the piston andfurther extends into a cavity of the slider bearing 108. In particularembodiments, the output shaft 112 and eccentric shaft pin 116 aremonolithic, i.e., made from one continuous piece of material, forexample by injection molding, casting, or by machining from a singlepiece of material.

In FIGS. 1 and 2 below the eccentric shaft pin 116 and assembled ontooutput shaft 112, there is a first radial bushing 120 and a secondradial bushing 122. The two annular bushings 120, 122 are arranged onthe output shaft 112 spaced some distance apart from each other alongthe output shaft 112. Between the two bushings 120, 122, there is afirst shaft seal 124 and a second shaft seal 126 also spaced apart fromeach other along the output shaft 112. Like the bushings 120, 122, inthe embodiment of FIGS. 1 and 2, the first shaft seal 124 and secondshaft seal 126 are annular.

FIG. 3 is a perspective cut-away view of the scotch yoke actuator 100(see FIGS. 1 and 2) that includes an enclosed cylinder 130 which housesthe power piston 102, and a shaft housing 132 that houses output shaft112. In certain embodiments of the invention, the cylinder 130 isconfigured to hold hydraulic fluid which is used to move the powerpiston 102 linearly in a reciprocating manner within the cylinder 130.As shown in FIG. 3, the output shaft 112 is supported within the shafthousing 132 by the first radial bushing 120 and the second radialbushing 122. The first radial bushing 120, second radial bushing 122,first shaft seal 124, and second shaft seal 126 are surrounded by anoutput shaft bearing 134.

FIGS. 4 and 5 show different perspective views of an air valve 200incorporating the scotch yoke actuator 100 of FIGS. 1-3 according to anembodiment of the invention. The air valve 200 includes a valve body 202and a valve member 204, the location of the valve member 204 isindicated, but not explicitly shown, in the illustrations of FIGS. 4 and5. However, one of ordinary skill in the art will be able to see how theair valve 200 functions. In a typical embodiment, the valve is a ballvalve and the valve member will be a ball of the type typically used ina ball valve, or the valve is a butterfly valve and the valve memberwill be a disk of the type typically used in a butterfly valve.

Movement of the valve member 204 is controlled by the scotch yokeactuator 100. An electrohydraulic servo valve 206 is coupled to thecylinder 130 such that the electrohydraulic servo valve 206 controls theposition of the power piston 102 within the cylinder 130 by controllinga flow of hydraulic fluid into and out of the cylinder 130. Inparticular embodiments, the air valve also includes a position sensor208, such as a linear variable differential transformer, to determinethe position of the power piston 102 within the cylinder 130.

In operation, the electrohydraulic servo valve 206 provides a flow ofhydraulic fluid into the cylinder 130 to move the piston linearly backand forth in a reciprocating manner within cylinder 130. As the powerpiston 102 moves back and forth, the eccentric pin 116 on output shaft112 rotates. This rotation is made possible by the slider bearing 108which slides back and forth in a direction transverse to the movement ofthe power piston 102. As can be seen from FIGS. 1 and 2, thesimultaneous linear movement of the power piston 102 and slider bearing108 allows the eccentric shaft pin 116 to move in a circular motion torotate the output shaft 112. The rotating output shaft 112 controls theposition of the valve member 204 either directly, as will be describedbelow and illustrated in FIGS. 6A and 6B, or indirectly by connecting tothe valve member 204 via a linkage mechanism as in the embodiment ofFIGS. 4 and 5.

FIGS. 6A and 6B show a plan view and top view for an alternateembodiment of an air valve 300 incorporating a scotch yoke actuator 302according to an embodiment of the invention. The scotch yoke actuator302 includes piston 304, slider bearing 306, and eccentric shaft pin308. The air valve 300 further includes electrohydraulic servo valve 310with position sensor 312, which may be a linear variable differentialtransformer.

In an exemplary embodiment, the air valve 300 of FIGS. 6A and 6B employsa butterfly valve, however, this configuration could be used with anumber of other types of valves. As such, the specific valve is notshown in FIG. 6A, as the claimed scotch yoke actuator 302 could operatea range of valve types. However, one of ordinary skill will readilyunderstand how the exemplary butterfly valve, for example with abutterfly plate disposed within a valve housing, would be incorporatedwith the scotch yoke actuator 302 shown. The scotch yoke actuator 302operates to rotate the eccentric shaft pin 308, which is coupled to anintegral output shaft 318, which, in the exemplary embodiment, directlycontrols the position of the butterfly plate to regulate the flow of airthrough the valve housing. A first shaft seal 320 and second shaft seal322 are assembled onto the integral output shaft 318. In the embodimentof FIGS. 6A and 6B, a first shaft bearing 324 is assembled onto theintegral output shaft 318 between the first and second shaft seals 320,322.

The embodiments of the scotch yoke actuator 100, 300 as described aboveand shown in FIGS. 1-6B address several of the problems associated withlinear crank sliders. As referenced above, some of the problems solvedby the present invention include elimination of the large overhangingmass typically associated with linear crank sliders. As such, the scotchyoke actuator of the present invention is lighter and more compact thanconvention actuators.

Linear crank sliders typically incorporate a linear seal with a scraperat the hydraulic fluid/air boundary. These seals become worn andpossibly damaged due to the repetitive wiping motion of the seal alongthe piston cylinder. Solid impurities in the hydraulic fluid canincrease the damage to these seals if the impurities get between theseal and the wiping surface. As such, the seals must be frequentlyreplaced or fluid leakage can occur. The annular seals 120, 122 used inthe present invention do not undergo the same wiping motion as sealsused in linear crank sliders, and the annular seals 120, 122 are notlocated at the hydraulic fluid/air boundary and are thus less likely toencounter solid impurities. Instead, the annular seals 120, 122 of thepresent invention are subjected to the rotation of the output shaft 112,which is less harsh than the wiping motion seen by linear crank sliderseals.

The monolithic, one-piece construction of the output shaft 112 andeccentric shaft pin 116 provides a more reliable, less expensivesolution than linear crank sliders with respect to the conversion oflinear motion to rotational motion. Also, the self-lubricating propertyof the slider bearing minimizes wear and increases reliability. Inparticular embodiments of the invention, a gimbaling connection, asdescribed above, allows for reliable operation of the actuator even whenthe shaft is deflected from its normal orthogonal position with respectto the piston and cylinder.

The perspective view of FIG. 7 shows an alternate embodiment of theinvention that includes a stepper-motor-driven scotch yoke actuator 400.The stepper-motor-driven scotch yoke actuator 400 eliminates the needfor a position sensor and position feedback. The hydraulic amplificationthat is typically provided, for example by an electrohydraulic servovalve (EHSV) flapper valve, is eliminated and replaced with a constantgain cam-nozzle amplification-tracking system. The combination of thecam-nozzle, stepper motor, and a gearbox in communication with the rackof the actuator piston provides an accurate and robust actuationpositioning system. One feature of this embodiment is that it canprovide a “fail-fixed” system, that is, a system that maintains its lastcommanded position in the event of electrical power failure.

In the embodiment of FIG. 7, the stepper-motor-driven scotch yokeactuator 400 includes a stepper motor 402. As will be shown below, thestepper motor 402 has an output shaft with a pinion gear coupled to aplurality of planetary gears, which are in turn coupled to a ring gear404. The ring gear 404 is coupled to a rack 406. An output shaft 408 ofthe planetary gears is coupled to a cam disposed inside of a firstcylinder 410. The cam is used to position a control piston, the workingsof which will be described in more detail below.

The rack 406 is disposed on a surface of a second cylinder 412 whichhouses the power piston 102 with cross-bore opening 106, and sliderbearing 108 (see FIGS. 1-3) for the stepper-motor-driven scotch yokeactuator 400. An output shaft 414 includes an eccentric pin 416 which isinserted through a slot in the second cylinder 412 and into the cavityof the slider bearing 108 (see FIGS. 1-3). As described in theembodiments above, the output shaft 414 rotates to control a valvemember for an air valve. Thus, in some embodiments, the bottom end 418of the output shaft 414 is coupled to a valve body 202 and a valvemember 204, as shown in FIGS. 4 and 5. In particular embodiments, thevalve may be a ball valve or butterfly valve, but is not limited tothese valve types.

FIGS. 8-14 provide a detailed view of exemplary embodiments of thestepper-motor-driven actuation system. With reference to FIGS. 8 to 10,a stepper motor 500 is used to drive cam 502. The stepper motor drives aplanetary gear system 504 where the ring gear 506 is in mesh relation torack 508. The pinion gear 510 is integral to the stepper motor rotor512. When the stepper motor 500 is rotated, the pinion gear 510 rotates.The pinion gear 510 rotation caused the planet gears 514 and planetframe 516 to rotate. The output shaft 518 is attached to the planetframe 516 and rotates with it. Similarly, the cam 502 that is attachedto the output shaft 518 rotates with the output shaft 518.

The cam rotation increases the gap between the cam 502 and nozzle 520 onone side of the cam 502 and decreases the gap between the cam 502 andnozzle 520 on the other side. The differences in the gaps affect the Pz1and Pz2 pressures on the ends 524 of the control piston 522 so as toforce the control piston 522 in the direction that will re-equalize thecam-nozzle gaps. The control piston translation opens the head port 526and rod port 528 to supply or drain, thereby causing the actuator piston530 and rack 508 to stroke. The rack 508 provides direct actuatorposition feedback to the ring gear 506, causing the ring gear 506 torotate. The ring gear rotation causes the planet gears 514 and planetframe 516 to rotate back to their original position, thereby rotatingthe cam and translating the control piston 522 to the mechanical nullposition (i.e., the center position).

When the cam 502 is in the center position, the hydraulic flow willenter port 534, pass through the cam-nozzle-orifice system (i.e., aroundcam 502 and through nozzles 520 and corresponding orifices), enter line536, and then drain out through Pb port 538 due to the lower pressure inthe Pb drain. It should be noted that the direction of flow is from line534 and into the nozzles 520 via the cam-nozzle gap (i.e., “flow in”) ascompared to conventional valves where flow is from the piston ends 524out of the nozzle 520 (i.e., “flow out”).

Note that when the cam 502 is positioned such that the control piston522 is towards the left-most position 540 in the control valve body 532,the supply port 534 is opened to the head port 526 (see FIG. 12). Whenthis occurs, the hydraulic flow passes through port 534, out throughhead port 526 and returns through rod port 528 and discharges outthrough Pb port 538. When the cam 502 is positioned such that thecontrol piston 522 is towards the right-most position 542 in the controlvalve body 532, the supply port 534 is opened to the rod port 528 (seeFIG. 11). The hydraulic flow passes through port 534, out through rodport 528 and returns through head port 526 and discharges out Pb port544.

During normal operation with a properly sized hydraulic andelectromechanical system, it is unlikely that the control piston 522will be at either its left-most position 540 or its right-most position542 (as respectively shown in FIGS. 12 and 11) due to the response ofthe system. In FIG. 11, the hydraulics are driving the actuator in theretract direction but it is against the retract stop. FIG. 12 depictsthe hydraulics driving the actuator in the extend direction but theactuator is against the extend stop.

Generally, as the stepper motor 500 rotates the cam 502, the controlpiston 522 begins to move and flow enters into either the head port 526or the rod port 528. As the control piston 522 continues to move due torotation of the cam 502, the port through which flow enters (i.e., headport 526 or rod port 528) opens wider, thus increasing the flow. As theflow pushes actuator piston 530, the rod 508 moves, thus rotating thering gear 506 as described above. The rotation of the ring gear 506 byrod 508 results in the cam 502 and control piston 522 translating to themechanical null position, thus preventing further flow to the actuator.The result is a proportional tracking of the actuator piston 530 to themotor rotor 512. As long as the dynamics of the system are sufficientfast so as to keep up with the input from the motor 500, the actuator530 will track the motor 500 commands with relatively small transientrotations of shaft 518, cam 502 and translations of control piston 522.

The primary disturbance to the system is the force input to theactuator. Any movement of the actuator piston 530 will cause rack 508translation and ring gear 506 rotation. Any ring gear movement resultsin cam 502 rotation due to the precision gearbox system 504. The highpressure gain of the system assures control piston 522 movement for anycam 502 rotation. The high pressure gain of the control valve ports 526& 528 coupled with the large head/rod areas will result in the requiredresistive force with minimal position error.

The stepper motor 500 and accompanying components described above couplea relatively low-energy motor with relatively high-energy hydraulics.Combining the stepper motor 500 with a suitable gearbox 504 provides thecapability to decrease stepper motor speed and increase its torque whilestaying at the same energy level. This is accomplished by properselection of the motor stator, rotor tooth count, and gearbox ratio.These components can be used to increase the motor torque, decrease it'ssusceptibility to torque disturbances, and still keep the motor fastenough to handle dynamic operation. Generally, the stepper motor 500 hasnearly perfect gain and is essentially unaffected by torque disturbancesdue to higher torque capability, gear box torque amplification, and theinherent detent feature of the stepper motor. In the embodiment shown,the round, symmetrical, balanced construction of the stepper motor 500is in essence unaffected by vibration and temperature variations.

The precision machined placement of stator and rotor teeth provide theinherent baseline position and gain accuracy of the stepper motor 500.If an accurate calibration is made, and the effects of disturbances arenegated, the need for a sensor is eliminated. This accuracy does notchange with life, is essentially constant from unit to unit, and is nota function of any calibration procedure. In certain embodiments, theround, symmetric construction of the stepper motor 500 maintains thisaccuracy in the presence of environment variations (e.g., temperature).Torque disturbances at the output shaft 518 such as dynamic sealfriction, nozzle hydraulic loads, unbalanced cam mass, etc. are minimaland are essentially rejected by the precision gearbox 504 (comprisingpinion gear 510, ring gear 506, planetary gears 514, and planet frame516) and the high detent torque of the stepper motor 500.

The detent torque prevents disturbances from having any appreciableeffect until they reach such a magnitude that they completely overpowerthe stepper motor 500. In particular embodiments, the stepper motorrotor rides on precision ball bearings and is inherentlyprecision-balanced about its rotation axis in the presence oftranslational accelerations (i.e., vibration), so the torquedisturbances at the motor rotor are negligible. The stepper motor 500has no channel-channel tracking error due to the fact that both channelsshare the same rotor-stator-pole flux circuit. Power transients have noeffect on steady-state operation, and the precision gearbox has minimalbacklash. In one embodiment, the backlash of the gearbox 504 isapproximately two step increments of the stepper motor 500.

With a thorough understanding of the two-nozzle embodiment firmly inhand, attention is now directed to FIGS. 13 and 14 which depictsingle-nozzle embodiments of the present invention. Each of thesesingle-nozzle embodiments operate similar to the two nozzle embodimentsdiscussed above and reduce cost over the two nozzle embodimentsdiscussed above, albeit at the expense of a reduced force gain and forcemargin. Each embodiment utilizes a means for applying force on one endof the control piston 522. Specifically, FIG. 13 illustrates anembodiment wherein the means is a spring preload with a constantpressure (Pb). FIG. 14 depicts a single-nozzle embodiment that does notinclude a spring preload, but instead utilizes a double-diameter endportion with a supply pressure on one diameter and pressure Pb on theother diameter.

With regard to the embodiment of FIG. 13, a spring 560 provides apreload on the control piston 522, e.g. a 10 lb. preload in oneembodiment. Those skilled in the art will recognize that other preloadforces may be provided depending on the operating parameters andconditions of the particular installation, and therefore all suchpreloads are to be included herein. This load is balanced by aPz1-induced force existing at the opposite end of the control piston522. Pz1 is regulated between the pressure at port 534 (Pc) and thepressure at port 544 (Pb) as a function of the cam 502 and the nozzlegap. In the presence of a constant spring force (spring scale isneglected) and constant Pc-Pb pressure, the fluid gap between the cam502 and the nozzle 520 is constant. This assures that the pistonposition is a function of cam position, and only cam position.

In the single-nozzle embodiment depicted in FIG. 14, the spring preloadis replaced by a hydraulic load via a double-diameter end portion 562.The double-diameter end portion 562 is desirable in that its hydraulicforce varies with Ps (at port 534)-Pb (at port 544) in the same mannerthat the Pz1 force does on the opposite end of the control piston 522.This trait ensures that the cam-nozzle gap stays constant in thepresence of a varying Ps-Pb, which ensures that the piston position 522is a function of the cam 502, and only the cam 502.

As can be seen from the foregoing, a robust stepper-motor-drivenproportional actuator has been described. Robustness, as used herein,refers to the ability of the system to remain accurate in the presenceof disturbance inputs and environment variations. Disturbances lead to ashift in the entire step-versus-position plot and gain variations leadto changes in the slope of the plot. Disturbances are due to undesiredtorques and forces as well as imperfect calibration. Gain variations aredue to the change in output/input characteristics due to component lifeand environment. Robustness is obtained by embodiments of the inventionby minimizing the magnitude of disturbances where possible, by isolatingthe device from disturbances where necessary, by maximizing thedisturbance rejection characteristics of the device, by designing adevice with minimal wear and/or whose performance is unaffected by wear,by precision calibration, and by inherent gain accuracy in the presenceof environment variations (e.g., temperature, stray flux, vibration,pressures, etc.).

All references, including publications, patent applications, and patentscited herein are hereby incorporated by reference to the same extent asif each reference were individually and specifically indicated to beincorporated by reference and were set forth in its entirety herein.

The use of the terms “a” and “an” and “the” and similar referents in thecontext of describing the invention (especially in the context of thefollowing claims) is to be construed to cover both the singular and theplural, unless otherwise indicated herein or clearly contradicted bycontext. The terms “comprising,” “having,” “including,” and “containing”are to be construed as open-ended terms (i.e., meaning “including, butnot limited to,”) unless otherwise noted. Recitation of ranges of valuesherein are merely intended to serve as a shorthand method of referringindividually to each separate value falling within the range, unlessotherwise indicated herein, and each separate value is incorporated intothe specification as if it were individually recited herein. All methodsdescribed herein can be performed in any suitable order unless otherwiseindicated herein or otherwise clearly contradicted by context. The useof any and all examples, or exemplary language (e.g., “such as”)provided herein, is intended merely to better illuminate the inventionand does not pose a limitation on the scope of the invention unlessotherwise claimed. No language in the specification should be construedas indicating any non-claimed element as essential to the practice ofthe invention.

Preferred embodiments of this invention are described herein, includingthe best mode known to the inventors for carrying out the invention.Variations of those preferred embodiments may become apparent to thoseof ordinary skill in the art upon reading the foregoing description. Theinventors expect skilled artisans to employ such variations asappropriate, and the inventors intend for the invention to be practicedotherwise than as specifically described herein. Accordingly, thisinvention includes all modifications and equivalents of the subjectmatter recited in the claims appended hereto as permitted by applicablelaw. Moreover, any combination of the above-described elements in allpossible variations thereof is encompassed by the invention unlessotherwise indicated herein or otherwise clearly contradicted by context.

What is claimed is:
 1. An actuator for converting linear motion intorotational motion, the actuator mechanism comprising: a cylinder havinga longitudinal axis and a power piston disposed in the cylinder, thepower piston configured to move in a reciprocating manner back and forthwithin the cylinder along the longitudinal axis; a slider bearingdisposed within a cross-bore opening in the power piston, the cross-boreopening having a slider bearing axis that is orthogonal to thelongitudinal axis, the slider bearing configured to move back and forthwithin the cross-bore opening; an output shaft having a central axisthat is orthogonal to both the longitudinal axis and the slider bearingaxis, the output shaft including an eccentric shaft pin inserted througha slot in the power piston and disposed within a cavity of the sliderbearing, the eccentric shaft pin being offset from the central axis ofthe output shaft; wherein the reciprocating movement of the power pistoncauses the slider bearing to move back and forth rotating the outputshaft.
 2. The actuator of claim 1, wherein the slider bearing is madefrom a self-lubricating material.
 3. The actuator of claim 1, furthercomprising a first radial bushing assembled onto the output shaft andadjacent an end of the output shaft which includes the eccentric shaftpin.
 4. The actuator of claim 3, wherein the first radial bushing is anannular radial bushing.
 5. The actuator of claim 3, further comprising afirst shaft seal assembled onto the output shaft to prevent a leakage offluid from the cylinder to the output shaft.
 6. The actuator of claim 5,further comprising a second radial bushing and second shaft seal eachassembled onto the output shaft proximate the first radial bushing andfirst shaft seal.
 7. The actuator of claim 5, wherein the first shaftseal is an annular seal.
 8. The actuator of claim 1, wherein the outputshaft and eccentric shaft pin are monolithic being constructed as asingle component.
 9. An actuator for converting linear motion intorotational motion, the actuator comprising: a power piston having alongitudinal axis, the power piston configured to move in areciprocating manner back and forth along the longitudinal axis; aslider bearing disposed within a cross-bore opening in the power piston,the cross-bore opening having a slider bearing axis that is orthogonalto the longitudinal axis, the slider bearing configured to move back andforth within the cross-bore opening; an output shaft having a centralaxis that is orthogonal to both the longitudinal axis and the sliderbearing axis, the output shaft including an eccentric shaft pin insertedthrough a slot in the power piston and disposed within a cavity of theslider bearing, the eccentric shaft pin being offset from the centralaxis of the output shaft; a stepper motor having a rotor; a camoperatively connected to the stepper motor, the cam position rotating inresponse to rotation of the rotor, the cam disposed within a controlvalve body having an inlet port, a rod port, a head port and at leastone drain port; a control piston positioned within the control valvebody and having a nozzle positioned on a first side of the cam in closeproximity to a surface of the cam and movable between a null positionand flow positions, the nozzle having a flow path leading from a firstend of the control piston, the control piston further including a meansfor applying a force on a second end of the control piston, the controlpiston moving as a result of a pressure imbalance at the ends of thecontrol piston occurring in response to a change in position of the cam;an actuator piston that is operably coupled to the cam, the actuatorpiston having a first side and a second side, the first side being influid communication with the head port and the second side being influid communication with the rod port, the cam moving in response tomovement of the actuator piston; wherein the actuator piston generatesthe reciprocating movement of the power piston which, in turn, causesthe slider bearing to move back and forth rotating the output shaft. 10.The actuator of claim 9, wherein the slider bearing is made from aself-lubricating material.
 11. The actuator of claim 9, wherein theoutput shaft and eccentric shaft pin are monolithic being constructed asa single component.
 12. The actuator of claim 9, wherein the cam isconnected to a gear shaft which rotates the cam, the actuator furthercomprising a gearbox connected between the rotor and the gear shaft. 13.The actuator of claim 9, wherein the translation of the control pistonopens one of the head port and rod port to supply and the other of thehead port and rod to drain, thereby causing the actuator piston and rackto stroke.
 14. The actuator of claim 13, wherein translation of thecontrol piston towards a first end of the control valve body opens thehead port to supply and the rod port to drain, and wherein translationof the control piston towards a second end of the control valve bodyopens the head port to drain and the rod port to supply.
 15. Theactuator of claim 13, wherein the actuator piston rotates the cam andtranslates the control piston to a mechanical null position in responseto the rack stroking.
 16. The actuator of claim 9, wherein the means forapplying a force on a second end of the control piston comprises aspring in operative contact with the second end of the control piston,wherein the second end of the control piston is in fluid communicationwith the drain port.
 17. The actuator of claim 9, wherein the means forapplying a force on a second end of the control piston comprises adouble diameter end portion having an end thereof in fluid communicationwith the inlet port, and wherein a hydraulic pressure at the end of thedouble diameter varies in the same manner as a hydraulic pressure at thefirst end of the control piston.
 18. An air valve comprising: a valvehousing; a moveable valve member disposed within the valve housing; avalve actuator coupled to the valve member, the valve actuatorcomprising: a power piston having a longitudinal axis, the power pistonconfigured to move in a reciprocating manner back and forth along thelongitudinal axis; a slider bearing disposed within a cross-bore openingin the power piston, the cross-bore opening having a slider bearing axisthat is orthogonal to the longitudinal axis, the slider bearingconfigured to move back and forth within the cross-bore opening; anoutput shaft having a central axis that is orthogonal to both thelongitudinal axis and the slider bearing axis, the output shaftincluding an eccentric shaft pin inserted through a slot in the powerpiston and disposed within a cavity of the slider bearing, the eccentricshaft pin being offset from the central axis of the output shaft;wherein the reciprocating movement of the power piston causes the sliderbearing to move back and forth rotating the output shaft.
 19. The airvalve of claim 18, wherein the output shaft extends through the valvehousing and controls a rotational movement of the valve member.
 20. Theair valve of claim 19, wherein the air valve is a butterfly valve, andthe valve member is a disk which the output shaft rotates to control aflow of fluid through the valve housing.
 21. The air valve of claim 19,wherein the air valve is a ball valve, and the valve member is a ballwhich the output shaft rotates to control a flow of fluid through thevalve housing.
 22. The air valve of claim 18, further comprising anelectrohydraulic servo valve coupled to a cylinder which houses thepower piston, the electrohydraulic servo valve configured to controllinear movement of the power piston.
 23. The air valve of claim 22,further comprising a linear variable differential transformer disposedin the cylinder, the linear variable differential transformer configuredto provide positional information for the power piston.
 24. The airvalve of claim 22, further comprising a first radial bushing assembledonto the output shaft and adjacent an end of the output shaft whichincludes the eccentric shaft pin.
 25. The air valve of claim 24, furthercomprising a first shaft seal assembled onto the output shaft to preventa leakage of fluid from the cylinder to the output shaft.
 26. The airvalve of claim 25, further comprising a second radial bushing and secondshaft seal each assembled onto the output shaft proximate the firstradial bushing and first shaft seal.
 27. The air valve of claim 18,wherein the slider bearing is made from a self-lubricating material. 28.The air valve of claim 18, wherein the output shaft and eccentric shaftpin are monolithic, being constructed as a single component.
 29. The airvalve of claim 18, further comprising: a stepper motor having a rotor; acam operatively connected to the stepper motor, the cam positionrotating in response to rotation of the rotor, the cam disposed within acontrol valve body having an inlet port, a rod port, a head port and atleast one drain port; a control piston positioned within the controlvalve body and having a nozzle positioned on a first side of the cam inclose proximity to a surface of the cam and movable between a nullposition and flow positions, the nozzle having a flow path leading froma first end of the control piston, the control piston further includinga means for applying a force on a second end of the control piston, thecontrol piston moving as a result of a pressure imbalance at the ends ofthe control piston occurring in response to a change in position of thecam; an actuator piston that is operably coupled to the cam, theactuator piston having a first side and a second side, the first sidebeing in fluid communication with the head port and the second sidebeing in fluid communication with the rod port, the cam moving inresponse to movement of the actuator piston; wherein the actuator pistongenerates the reciprocating movement of the power piston which, in turn,causes the slider bearing to move back and forth rotating the outputshaft.
 30. The actuator of claim 29, wherein the cam is connected to agear shaft which rotates the cam, the actuator further comprising agearbox connected between the rotor and the gear shaft.
 31. The actuatorof claim 29, wherein the translation of the control piston opens one ofthe head port and rod port to supply and the other of the head port androd to drain, thereby causing the actuator piston and rack to stroke,wherein translation of the control piston towards a first end of thecontrol valve body opens the head port to supply and the rod port todrain, and wherein translation of the control piston towards a secondend of the control valve body opens the head port to drain and the rodport to supply, and wherein the actuator piston rotates the cam andtranslates the control piston to a mechanical null position in responseto the rack stroking.
 32. The actuator of claim 29, wherein the meansfor applying a force on a second end of the control piston comprises oneof: a spring in operative contact with the second end of the controlpiston, wherein the second end of the control piston is in fluidcommunication with the drain port; and a double diameter end portionhaving an end thereof in fluid communication with the inlet port, andwherein a hydraulic pressure at the end of the double diameter varies inthe same manner as a hydraulic pressure at the first end of the controlpiston.